The invention relates to a parallel-and-external-axis screw-type rotary piston compressor comprising a housing provided with inlet and discharge ports and at least two rotors as the main and gate rotors, which are edged with helically extended tooth spaces and are arranged to interengage and with their axes parallel to each other. The tooth profiles of the main rotor are designed to be substantially convex and outside of the pitch circle and the tooth flanks of the gate rotor being designed to be substantially concave and within the pitch circle.
Although screw-type compressors have been built for about thirty years, it was only in the past fifteen years that they have been widely used as series machines. In terms of quantities, it is primarily the smaller air or refrigerant compressors cooled by injected oil that constitute the major portion of the world production of screw-type compressors.
Screw-type compressors mostly are two-shaft rotary piston compressors. They operate in a similar manner as the known piston compressors, according to the principle of displacement. In the case of screw-type compressors, the operation spaces are constituted by the tooth spaces of two interengaging rotors with helical gears, which are revolving in a housing closely surrounding the rotors.
On rotating the compressor rotor, which is provided with a special toothing, the tooth space volume of a pair of tooth spaces that exists between the end face of the rotor and the lines of contact of the rotor teeth gradually declines from a maximum value to zero. The gas to be conveyed, which is trapped between the tooth spaces and the bores of the housing, thus is constantly compressed and is finally displaced out to the pressure pipe through a dislodge port provided in the compressor housing. Since there are no mass forces, screw-type compressors can operate at high speeds and thus can be constructed to be small and light.
A substantial disadvantage of screw-type compressors is the imperfect sealing of the operation spaces between the tooth crests and the compressor housing, between the rotor end faces and the end lids of the housing as well as at the rotor mesh. Due to the internal leakage, that is, the amount of gas flowing off through the leakage gaps of the operation space, the efficiency of the compressor is considerably deteriorated, the loss of effeciency depending on the shape of the toothing chosen as well as on the production accuracy achieved.
In order to keep these internal leakage amounts small, hydraulic oil is injected into the tooth spaces during the compression procedure with most screw-type compressors, the amount of oil injected being about 6 to 15 l/m.sup.3 of intake air. The oil is most finely dispersed because of the high rotor speeds, thus forming a two-phase mixture with the air, whereby it is possible to reach a better sealing of the leakage gaps and thus a reduction of the leakage-gas amounts. On the other hand, the addition of oil results in increases internal losses within the compressor, because of oil specific effects such as splash losses and incompressibility adversely affects the efficiency of the compressor and restricts the amount of oil per intake-air amount.
A further significant disadvantage of the screw-type compressors available on the market results from the complexity of the special toothing used, by which it is aimed at reaching particularly large compression chambers with leakage gaps as small as possible. A tooth flank profile of this type is known, for instance, from German Offenlegungsschriften Nos. 26 39 870 and 27 35 670. There, tooth flank profiles for helical rotors are described, which, in the end cross section, are comprised of a plurality of flank parts, such as circular arcs, elliptical curves, involutes, cycloids and hyperbolic curves. Because of the complexity of these flank curves, extremely sophisticated production methods and hence very expensive tools are usually required, which often do not allow for an economical production of the rotors, in particular for small screw-type compressor units. In addition to the numerous operation steps and the various tools necessary to carry out the same, the realization of such complex profiles frequently is adversely affected by unfavorable cutting conditions so that the tools wear to an increased extent, shortening their service lives. Frequent re-grinding of the tools is necessary, which results not only in increased costs for the operation procedure as such, but also in reduced accuracy of the profile form and thus frequent calls for expensive refinishing operations.
In German Auslegeschrift No. 22 34 777 a tooth profile is described, which is comprised of involutes and circular arcs. With this arrangement, a further involute is generated on the gate rotor as the envelope of the main rotor involute associated with the cycloid. Since this involute extends beyond the pitch circle of the toothing because of the profile form of the main rotor, a relatively long involute section is generated on the gate rotor, which causes a large blow hole. The blow hole of the compressor toothing is formed because the rotor mesh along which the tooth flanks of the main and gate rotor teeth of an interengaging tooth pair meet does not extend as far as to that edge of the housing which results from the intersection of the two housing bores.
Furthermore, a relative velocity is produced on the gate-rotor head, that depends on the radius of the gate rotor, if this radius does not coincide with the radius of the pitch circle as appears from to German Auslegeschrift No. 22 34 777. This relative velocity causes scuffing and increases the leakage amounts in addition to the large blow hole, thus deteriorating the internal compressor efficiency.
In German Offenlegungsschrift No. 31 40 107 a rotor profile is proposed, in which the tooth flanks of the main rotor are not composed of curve segments, but are formed by a continuous, uniform analytically definable curve form from one crest point of the main rotor to the next one. Although this results in a very simple and robust design of the main rotor, which can be produced in a relatively simple manner, the form of the toothing of the gate rotor is markedly more complicated, which again involves production problems. Furthermore, this toothing, due to the profile form, has very small operation spaces with comparatively large clearance lengths. Despite the fact that the production of the main rotor has been simplified and thereby the toothing can be produced more accurately, relatively large leakage gaps form, which produce large leakage-gas amounts and thus lower compressor efficiencies than comparable systems.
In addition, intensive scuffing and heating during operation are to be expected, in particular with the embodiment known from German Offenlegungsschrift No. 31 40 107, which has pointed teeth on the gate rotor, whereby the leakage-gas amounts are again raised and the compressor efficiency is again lowered.